DE4126650A1 - Vehicle gear change device - has two change couplings with intermeshing coupling rings, one with axial spring location - Google Patents

Abstract

The gear change device comprises at least two change couplings. At least one is a gear coupling, having a toothed coupling ring (24) such that the coupling moves axially via a switching device against an opposing coupling ring (23), both coupling rings intermeshing. The opposing coupling ring has spring location allowing axial movement. The gear change occurs in the synchronous region of the gear member to be coupled. ADVANTAGE - Guarantees smooth gear changes without affecting load characteristics.

Description

The invention relates to a switching device, in particular for automatic switching of multiple driving ranges with stepless Geared or multi-stage gearboxes according to one in the generic terms of claims 1 to 3 and 11 listed Art. Especially in the case of continuously variable branching gearboxes with positive or positive plus non-positive coupling units is a special type of control the range clutches necessary to achieve high shift quality without To achieve load interruption. Even when using a friction clutch in the form of a cone coupling, it is advantageous to use a targeted, adapted shifting sequence with appropriate hydrostat control to be realized within the switching phase.

In the case of continuously variable hydrostatic-mechanical branching transmissions, it is common for the hydrostatic units A and B to interchange in their function as a pump and motor after each range change or after each range shift. The load-dependent speed slip of the hydrostatic transmission or the shaft 5 connected to the hydrostatic unit B has the reverse effect after each range shift, which must be corrected by the adjusting device within the shift phase in order to ensure a pressure-free shift.

The object of the invention is a switching device, preferred for automatically switchable transmissions, especially for stepless branching gear with purely positive or positive-locking clutches that look like they are made of DE 41 04 167.4 known, or friction clutches with cone friction surfaces to create the high switching quality without interrupting the load.

This problem is solved by the in the main claims 1 to 5, 20th and 25 listed features solved. Further advantageous configurations go from the dependent claims and the description forth. The invention is explained with reference to drawings. It shows

Fig. 1 is a circuit diagram,

FIG. 1a is a further circuit diagram,

Fig. 2, the displacement and speed characteristics over a shift sequence,

FIG. 2a pressure curve of the hydrostatic transmission and torque characteristic of two couplings within a switching sequence,

Fig. 2b circuit diagram of a hydrostatic-mechanical device part based on the opening operation of a switching range clutch,

Fig. 3 is a form-plus-positive or purely form-fitting coupling device as a double clutch,

Fig. 3a shows an embodiment of the coupling profile,

FIG. 3b shows a coupling version in which the coupling ring and the piston constitute separate components,

Fig. 3c shows a further embodiment of a coupling with a hydraulic return of the pressure piston,

Fig. 3d and 3e shows a further embodiment of the coupling profile,

FIGS. 4 and 5 a shift clutch with synchronizing means,

FIGS. 6 and 6a, a coupling design as a friction clutch having conical friction surfaces, designed as a double clutch,

Fig. 6b is a friction clutch with additional shift toothing,

Fig. 7 to 7c different embodiments of the coupling profile with an axial torque-dependent force component by inclined driving surfaces.

In Fig. 1, the hydrostatic transmission is marked with 1 , the mechanical transmission part, which contains the range couplings K1, K2, K3 with the control lines 11, 12, 13 , with 2 and the control unit or the control logic with 6 . The hydrostatic transmission 1 and the mechanical transmission part are, for example, via planetary gear units, shafts and gears, e.g. B. as shown in Fig. 8, interconnected.

The leakage oil losses of the hydrostatic transmission cause an inevitable speed slip of a hydrostatic shaft 5 . This speed slip has an effect in connection with a continuously variable power split transmission with several driving ranges within the range shift, as shown in Fig. 2, so that, for example, in train operation to achieve the synchronous state of the clutch elements to be switched, the adjustment quantity by the dimension X at the shift point is larger must be the theoretical value in order to reach the synchronizing point of the coupling elements to be switched at switching point S. Due to the above-mentioned function reversal from pump to motor and vice versa after switching to the next switching range, the speed slip also has the opposite effect, which corresponds to an adjustment correction quantity by the dimension Y. As a rule, the correction values X and Y have different sizes, due to the respective transmission design and area sizes corresponding to the different sizes of the hydrostatic pressures p1 and Δp2 before and after the shift. During the shift phase, the hydrostatic transmission must be corrected in accordance with the load state by the correction dimensions X and Y. According to a switching example as shown in FIG. 2, where B1 means 1 and B2 means 2 , the functional sequence is as follows:

In the case of an upshift under load from B1 to B2, the switching signal for closing the second range clutch is triggered as soon as the shaft 5 or the coupling members connected to it with the coupling members to be shifted have reached synchronous operation at the switching point S. The hydrostatic adjustment is larger by X than the theoretical value. Now the shift into the second range or the closing of the second range clutch takes place after the synchronization pulse, the first range clutch remaining closed. The hydrostatic adjustment is now reduced by the dimension Z within both closed clutches until the setting point SK1 is reached, after which the signal for opening the first range clutch is issued. Only then is the gear ratio changed further with further hydrostatic adjustment in order to pass through the second area.

In view of an optimally short switching time, the hydrostatic adjustment device immediately after the second closed Area clutch reversed in the opposite direction. This can programmed electronically in the control logic or by a corresponding shuttle valve, not shown, for example depending on the control pressure pK2 for the second range clutch is reversed, can be realized.

The coupling shown in Fig. 3 corresponds to the coupling design as known from DE 39 03 010 and described in the preamble of claim 1. It has a clutch toothing 27 which has axial deflection surfaces, as shown in FIGS. 7 to 7c, so that an axial force component for disengaging the clutch is generated as a function of the torque as soon as the clutch 21 or 25 is depressurized by the disengagement signal or opening signal is set. In the switched-on state, the transmission member 19 is connected via the coupling ring 23 and the coupling toothing 27 to the coupling member 24 , which is non-rotatable and axially displaceable with the coupling carrier 20 and which at the same time represents the pressure piston under pressure, with the coupling carrier 20 , which at the same time forms the output shaft. The clutch 22 is preferably here, for. B. as a second range clutch, opened and thus no load.

The coupling profile 27 is also very advantageous according to the invention as a form-fitting coupling toothing, as shown in Fig. 3a 3e, 3d and described in more detail later. The clutch carrier described in claim 1 can generally be implemented as a rotating component or non-rotating component when a gear part (for example, a ring gear of a planetary gear) is connected to the housing (not shown).

According to the invention, a pressure valve 16; 122 is provided, which is reached after reaching the synchronous state on the new clutch via a synchronization signal e or a signal dependent thereon, so that the control pressure for the new clutch to be shifted is increased spontaneously in order to close this clutch as quickly as possible. This pressure build-up, which can also act in the form of a pressure surge, also enables the clutch to be closed despite certain synchronous deviations or the build-up of torque during the clutch closing process.

Furthermore, the invention provides a check valve 17 , which the inflow to the servo 4; 18 locks or blocks the adjustment device of the stepless converter after reaching the synchronous state of the new clutch until the new clutch is closed. This blocking valve 17 or this blocking device can also be controlled by the synchronizing signal e, as can the corresponding clutch valve 161; 162, 163, 164 of the new clutch ( Fig. 1a). After the synchronous operation has been achieved and the switching pulse has been initiated, this blocking device 17 prevents the synchronous state from possibly being left before the new clutch is closed. After the clutch is closed, the locking function of the adjusting device is immediately released by releasing the locking valve 17 so that the adjustment correction (x, y; z) already described can be initiated and completed within both closed clutches. The signal to cancel the blocking function can be triggered by the pressure signal 157; 158; 159 from the building up control pressure of the new clutch. A programmed lock-up time or depending on other company sizes can also be sufficient in some applications. The operating variables can be engine speed, temperature, accelerator pedal position, throttle valve position, hydrostatic pressure, rate of change of the adjustment and, above all, the delivery rate of the feed pump. In the size of the blocking time, the filling time of the pressure pistons of the new clutch has to be taken into account, which can vary depending on one or more of the aforementioned operating variables. The locking device for the adjustment can preferably also be programmed in the electronics or in a hydraulic and / or mechanical device, for. B. a known damping valve, the signals described above can serve as control signals. The locking function is implemented here by a corresponding valve 170 , which regulates the hydrostat adjustment. A shuttle valve 44 , which is connected between the inflow lines 116, 117 for hydrostatic adjustment and is controlled via control signals of the old and new clutch or the synchronizing signal e, is also suitable as a check valve (not shown).

The Druckvnetil 16; 122 can be a two-stage valve, ie with two pressure stages, the lower pressure stage for normal operation and the high pressure stage for switching the new clutch being activated briefly. The synchronizing signal e can also be used to control the second pressure stage at the same time.

To optimize efficiency, the first pressure stage is preferably load-dependent in a further exemplary embodiment, for. B. can be changed or modulated by the working pressure Δp1 or Δp2, as will be described in more detail elsewhere. According to the invention, an alternatively provided hydraulic accumulator 45 , which is always filled with the higher pressure of the second pressure stage after a switching operation, ensures sufficient oil quantity for the second pressure stage. A valve 46 closes the outlet of the hydraulic accumulator 45 so that the pressure is maintained until the next switching. When the new clutch is activated, this valve is also activated and opened. In the event of a pressure drop outside the circuit, the hydraulic accumulator is refilled by the main pressure valve 16; 122 is controlled to the second or high pressure level, whereby the hydraulic accumulator 45 is filled valve 46 open at the same time. This hydraulic accumulator 45 is preferably applicable for transmissions with load-dependent and / or speed-dependent pressure modulation or / and for transmissions with a small feed or supply pump for reasons of efficiency. The hydraulic accumulator 45 can also serve to increase the adjustment speed, in particular to meet the high reset requirement, e.g. B. in a car transmission in an extreme acceleration process or also to perform the braking process. The hydraulic accumulator is always switched on with the known means if the feed pump cannot maintain an adequate supply in order to meet the short-term peak requirements mentioned.

In an electronic control device such. B. a potentiometer signal the pressure variable Δp and adjust the adjustment angle α to the pressure-dependent leakage oil loss or the speed slip and accordingly shift the switching point S _K2 relative to the theoretical switching point S by the corresponding correction dimension X.

In a further embodiment for the shift correction variable, a correction mechanism acting on a shuttle valve or check valve 17 can also be used, which, depending on the respective load state, determines the switch-on point for the follow-up clutch in a load-dependent manner, eg. B. by the hydrostatic pressure signal and a position signal of the hydrostatic adjustment. The blocking position spontaneously blocks the hydrostatic adjustment until the new clutch is closed.

With an electrical or electronic hydrostat adjustment, e.g. B. by a potentiometer, the adjustment voltage can be overridden or superimposed by the load-dependent correction voltage. The effective adjustment voltage quantity corresponds to the adjustment quantity of the hydrostat or the adjustment angle α, which is corrected in the switching phase as a function of the load by the aforementioned correction variables X and Y, as shown in FIG. 2.

To achieve the synchronous state of the clutch elements to be shifted in the shift point, a speed comparison used in the control logic of two or more transmission elements, for example the comparison of an input speed signal a with an output speed signal c, can be used to generate the synchronous signal and the associated control signal for closing the corresponding clutch To signal K1 or K2 or K3. An adjustment signal d connected to an adjustment member is also suitable as a control signal for clutch actuation, a corresponding adjustment variable signaling the synchronization point of the clutch members concerned and depending on, for example, a load-dependent signal from working pressure 7; Δp1 of the hydrostatic transmission 1, an adjustment correction corresponding to the speed slip in the control unit or in the control logic is taken into account. The pressure signal Δp1 causes a correction of the adjusting element due to the leakage oil.

In Fig. 2, the characteristic behavior of the hydrostatic adjustment - adjusting unit A - in relation to the hydrostatic speed n _{B of} the hydrostatic unit B in the load state, z. B. shown as a train upshift over a range change B1 / B2.

The correction of the adjustment variable X of the adjustment angle α results from a load condition which occurs, for example, when the train is shifted up from a first area B1 to a second area B2. This correction variable can be processed electronically or hydraulically and mechanically, in that the pressure Δp1 of the working pressure line in the control logic signals a correspondingly larger adjustment angle until the switching signal is triggered to close the second range clutch K2. If this adjustment angle correction is implemented hydraulically or hydraulically and mechanically, a corresponding adjustment device 15 is used for this purpose, which preferably presses the working pressure Δp1 or Δp2 against a spring-elastic element and, depending on the corresponding spring travel, one relative to the theoretical adjustment angle αth Adjustment angle increased by dimension X until the switching signal for closing the second range clutch is triggered. The working pressures .DELTA.p1 and .DELTA.p2 have mutually opposite effects on the direction of adjustment of the adjusting member 14 , whereby the respective operating behavior - pulling or pushing operation - is taken into account. Dynamic operating changes immediately before or within the switching phase are hereby adapted in a targeted and spontaneous manner. The coupling designs provided allow slight deviations in synchronism without any noticeable influence on the shift quality. This can e.g. B. Deviations that result from different oil temperatures, oil viscosity, manufacturing tolerances and / or wear. The invention further provides a synchronizing device, as shown in FIG. 4, for all embodiments of the couplings with a coupling profile , in order to compensate for synchronous errors and to meet all requirements with regard to comfort, safety and stability.

The rate of change of the hydrostatic adjustment or the gear ratio also has an influence on the shift quality. A corresponding influencing variable can be generated, for example, by an engine operating variable, such as throttle valve position, accelerator pedal travel, magnitude of a braking force, as a control signal via a control line 57 on the shift control. The switch-off point Sk1 can be varied as a function of these aforementioned operating values, which can have advantageous effects depending on the application or type of motor vehicle.

Fig. 2a shows the pressure profile and .DELTA.P1 .DELTA.P2 and the torque behavior Tk1 and Tk2 within the switching time ts. The torque Tk2 of the second range clutch K2 increases continuously as soon as it is closed, the torque Tk1 of the first range clutch K1 automatically dropping, which is triggered by the hydrostatic position by the variable Z within both closed clutches K1 and K2. The working pressure Δp1 drops to a greater extent than the clutch torque TK1 and, with an intermediate manipulated variable "S", has an unpressurized state in which the full power is purely mechanical or without the involvement of the continuously variable converter via both closed clutches K1 and K2 or K2 and K3 is transmitted. At this point, Δp2 begins to rise.

FIG. 2b shows a control device which determines the favorable opening point of the range clutch K1 switched in the previous switching range B1 via a hydraulic or hydraulic-mechanical device. The device consists of a clutch switching valve 51 and a further valve 52 , which releases the hydrostatic pressure Δp1 within the first part of the switching phase via a control line 53 in order to protect the clutch switching valve 51 from opening prematurely.

With regard to efficiency optimization, the hydrostatic transmission 1 is provided with a variable feed pressure 10 which varies as a function of the working pressure Δp1 or Δp2. The feed pressure 10 also serves as the clutch pressure for the range clutches K1, K2, K3 etc. The control pressure 56 for the valve 521 described above, which determines the opening time of the previously closed clutch, is correspondingly variable. With correspondingly low load conditions, the control valve 52 is accordingly programmed to a very early opening point due to the constant spring force of the valve spring 55 , so that a quick switching sequence can take place despite low clutch pressures at low load conditions.

The functional sequence is as follows: after synchronous operation of the second range clutch or the follow-up clutch K2 to be switched, the opening signal is immediately issued via a control line 11 a. However, since the clutch K1 is not yet allowed to open, the hydrostatic pressure Δp1 is held in the switching position via the control line 53 and the clutch valve 51 until Δp1 has dropped to a favorable switch-off point, as a result of which the valve 52 is released in the control line by a decrease in the hydrostatic pressure Δp1 54 and is brought into the switched-off position by means of the compression spring 55 and thus the pressure in the control line 53 also drops, as a result of which the clutch switching valve 51 can go into the open position. The favorable switch-off point can be determined via a valve spring 55 , which is dependent on the type of clutch and the shift path of the clutch. Instead of the working pressure .DELTA.p1, the working pressure .DELTA.p2 of the other pressure side 8 of the hydrostatic circuit can also be used here, in that the clutch switching valve 51 can only move to the open position after the action of .DELTA.p2.

To optimize the range switching, as a further feature of the invention, especially in the case of electronic control devices, the hydrostatic pressure value Δp1 effective before the start of the shift saved and then the opening point for the first clutch K1 set, e.g. B. by a correspondingly large pressure value of Δp2. This can also be solved hydraulically and mechanically, by a control element depending on the load or pressure of Δp1 a control path variable is shifted to a control point that after an angular position α of an actuator of the hydrostat at corresponding pressure Δp the opening signal for the Clutch triggers.

Instead of the switching point-determining hydrostatic pressure values Δp1 or / and Δp2, a control pressure modulated by the hydrostatic pressure Δp, e.g. B. the feed pressure 10 can be used.

In order to prevent a shift jerk, it is important, as already described, that the opening signal is triggered after a hydrostatic reset within a correction variable "Z", namely between the setting points "S" and "SK1". This signal occurs according to the invention from a pressure signal Δp1 in the hydrostatic pressure side 7 of the hydrostatic circuit in such a way that after the second range clutch K2 has been switched on and reset by a correction dimension “X”, the hydrostatic pressure is relieved and the pressure Δp1 has almost dropped to zero, which is the adjustment point “S” corresponds. If the hydrostatic adjustment or the adjustment angle α is further reduced within the adjustment variable "Y", the clutch can open after the opening signal has been given as a result of the pressure drop Δp1. When using a positive and non-positive coupling, as described at the beginning and shown in FIGS. 3, 4 and 5, a safe opening process is ensured with simultaneous hydrostat adjustment, since this coupling can be pressed on depending on the torque. This has the advantage that a very fast and smooth shifting sequence is ensured in all operating situations, both in overrun or pull and upshift or downshift and in all load situations. In view of the fact that the area clutch to be opened is subject to a certain switching time or opening time, it makes sense to advance the opening signal somewhat. This can be influenced in such a way that the falling pressure line .DELTA.p1 can be set or programmed to a correspondingly larger pressure value. The positive and non-positive coupling, as shown in FIG. 3, offers the advantage here that after the opening signal is triggered and the control pressure drops thereafter in the corresponding first range clutch 21, the pressure piston 24 is pushed back via the torque-dependent deflection toothing or toothing slope in neutral position with additional restoring force of the spring 29 . Depending on the design of the clutch teeth 27 and the given switch-off path, the opening signal can be set at correspondingly different pressure sizes. The pressure signal .DELTA.p2 of the other pressure side 8 of the hydrostatic circuit can also be used to trigger the opening signal, with the previously closed area clutch K1 being actuated to open and brought into a depressurized state by the pressure increase in .DELTA.p2 after the middle adjustment point "S" has been reached. In this embodiment, the opening signal is triggered in the opposite manner to the previously described embodiment by building up pressure in one of the working pressure lines of the hydrostatic transmission. This type of control can be very advantageous for certain applications, e.g. B. be used in work machines such as tractors to improve efficiency if this point is placed in the main operating area, since in this transmission point the hydrostat is relieved and the hydrostat loss values are correspondingly low.

The first exemplary embodiment for forming the clutch opening signal has the advantage of being largely independent of the respective load state, since the pressure drop from Δp1 to zero forms a defined, relatively early opening time and, when using a positive-locking clutch, a certain opening time for disengaging the disengaging clutch after the opening signal has been given with the aid of the decreasing clutch torque of this clutch. When using the second alternative embodiment with the aid of the working pressure Δp2 of the other pressure side 8 , the disengaging process takes place after the switching point S, as a result of which a shorter disengagement time is adapted in accordance with a shorter disengagement path of the clutch to be opened. Here, a simple adaptation, in particular to the embodiment or the tooth height of the switching toothing 27 , FIG. 4, is possible.

In order to ensure that the old clutch is reliably switched off in all operating situations, in particular in the case of clutch designs which do not generate any axial force component in the closed state, as shown in FIG. 3a, the invention provides that in the switch position S _K1 a brief adjustment lock is established or is programmed to allow a certain opening time within this period in the no-load state of the old clutch in order to prevent a build-up of torque within an opening time which corresponds to an opening travel T. This adjustment lock can be implemented via the shut-off valve 17 or the main control valve 180 or by a permanently programmed function in the electronics or the control logic. In many applications, very large tooth play SV can advantageously be selected. The signal for triggering this function can expediently be triggered via the pressure signals .DELTA.p1 and .DELTA.p2 or a torque signal on one of the transmission elements, which signals the no-load condition of the old clutch.

The same switching functions apply to overrun mode, but with the difference that the correction of the calibration variables behave the other way round. This also applies, for example in an upshift or downshift.

The correction of the adjustment α or the adjustment variables is effected in these operating cases by the pressure Δp2 of the other pressure side 8 .

For vehicles with fixed translation settings, such as B. required for work machines, the invention provides a device in which the translation point is held at two range boundaries for continuous operation or for operating conditions with a longer dwell time at this translation point. The device already described can be used for this purpose, the old clutch remaining closed after the new clutch being closed and the adjusting device of the transmission or the adjusting angle α of the hydrostatic transmission being able to move within the correction range Y and X or Z depending on the load state. With regard to efficiency optimization, the invention alternatively provides a bypass valve, not shown, between the two working pressure lines 7 and 8 of the hydrostatic transmission, which can be opened in this operating state or transmission state with the two clutches closed in order to reduce hydrostatic losses. Before the old clutch or one of the two clutches is opened, the bypass valve is closed again depending on one or more operating variables in order to activate the hydrostatic pressure circuit. The closing process of the bypass valve is expediently carried out as a function of the adjustment angle α of the continuously variable transmission part, with the bypass valve closing process preferably being gradual, for. B. via appropriate throttle devices to ensure a continuous shock-free state. The bypass valve can be designed in various ways, for. B. also in the form that a certain differential pressure between the two working pressure lines 7 and 8 is maintained and, depending on the size of the correction value X or Y, has a different pressure size Δp. This differential pressure .DELTA.p can also serve as a control pressure or as a signal for closing the bypass valve, thereby signaling whether the transmission point should be left with both closed range clutches. Based on this signal size, the signal for opening one of the two clutches can then also be given. Which of the two clutches is opened depends on the desired or newly operated area B1 or B2. In the case of a version without a bypass valve, the opening signal for the old range coupling can, as already described, one or both of the operating pressures Δp1; Δp2 or / and Δp are used, which signal from their respective pressure values whether the state should be exited with both closed clutches or this transmission state. The change in angle of the adjustment angle α occurring within the correction range X, Y or Z or the change in the accelerator pedal or other operating variables which are suitable for causing or signaling a change in gear ratio can also be used as the opening signal of the old clutch. A fixed ratio setting within both closed clutches K1 and K2; K2, 3 can also be specified via a corresponding selection device 108 , which can be specified via a mechanical or / and electrical or / and hydraulic device.

In vehicles, e.g. B. Cars that allow short-term load interruptions, the bypass valve between the working pressure lines 7 and 8 can also be used advantageously, with the bypass valve being opened with simultaneous automatic gas reduction in the switching phase, for example, to synchronize the clutch version 21 , Fig. 21 or version 80 To minimize Fig. 6 with friction elements.

All signals required for the switching functions can be transmitted hydraulically and / or mechanically, electrically or electronically. As is known per se, potentiometers can be used very advantageously for signal variables for the adjustment angle, pressures, throttle valve position and accelerator pedal travel. An example of hydraulic switching function is shown in Fig. 2b, and described in more detail elsewhere.

A switching hysteresis device provided according to the invention ensures that the translation is switched on and off too frequently is prevented in the switching range. This facility can be controlled by a signal that results from a certain time variable and / or a difference in the load variable, e.g. B. working pressure Δp1 and Δp2 or / and an adjustment variable / Adjustment angle α or / and rate of change a company size, e.g. B. speed changes, accelerator-path changes or / and other company sizes. Hydraulically is z. B. a switch back to the previous area long prevented until a certain piston or a certain Valve member in a throttle valve a fixed limited Has traveled distance (not shown). At An electronic control device is the information for the switching hysteresis from the aforementioned operating variables in entered the microprocessor and processed accordingly.  

According to FIG. 3a shows an embodiment 71 of the coupling profile 27 which is not repellent cam surfaces 74 has, which are associated camming repellent 72nd The non-repelling driving surfaces 74 with a supporting height T have no or only a very small angle of inclination γ, which lies within the friction angle of all coupling elements. Under certain circumstances, this angle γ can also be designed to be negative, so that here the coupling rings are pulled together in the closed state. The return springs 28; 29 are expediently dimensioned in this embodiment such that the coupling members can be disengaged even in the load state after the switching pressure has been removed. This also applies to the embodiment according to FIG. 3 for the hydraulic return, the correspondingly high restoring forces in the pistons or piston surfaces 217; 219 .

The helix angle γb or γa of the inclined surfaces 72 (or 404 Fig. 3a, 3d) and the piston surfaces of the piston 24; 25 and the switching pressure are designed according to the invention so that in the switching phase over these repelling driving surfaces 72; 404 close the new clutch against the maximum clutch torque, ie the gear teeth 27; 401 can interlock. For this purpose, the switching pressure can, according to the invention, briefly interlock the switching teeth 27; 401 ; . . . be raised until the clutch is closed. This also applies to the coupling design with deflection teeth, which in the closed state has a torque-dependent axial force component, the axial deflection forces resulting from the inclined surfaces 522; 521; 313; 318 at an angle γ (see FIGS. 7 to 7c) is overcome by sufficient piston switching force.

When using the clutch toothing 71 , which in the closed state does not generate an axial torque-dependent force component, an opening process is only possible if the clutch is almost torque-free. This means that in the switching phase after the new clutch is closed, the old clutch only opens after the torque has dropped so far until the spring force of the return spring is sufficient to overcome the frictional resistance in the clutch teeth and the driving surfaces.

The load-bearing tooth height T of the coupling profile can be very high low level of e.g. B. less than 1 mm, depending on the size of the Torque and the structural conditions, so that the opening path can be kept very short. The tooth game SV can be made relatively large in most applications be what the turn-on process as well as the turn-off process the clutch favors.

The deflecting surfaces 72 also serve for the rapid decoupling and pushing back of the coupling ring in neutral position after the first opening travel according to the dimension T via the spring force of the return spring 28; 29 or a hydraulic return via pressure piston 217; 219 has taken place.

The coupling profile 71 has the advantage, among other things, that the backlash SV can be kept very small and nevertheless there is a large backlash at the start of engagement within the repelling driving surfaces 72 of both coupling rings, in order to ensure a secure interlocking of the coupling teeth.

As mentioned, the angle γ can also be designed in the negative direction - not shown - so that the coupling profile is pulled together depending on the torque. It is also possible to undercut the clutch toothing, as can be seen in FIG. 3d, so that hooking against release in the torque state is specifically achieved. It is thus possible that, after the opening signal has been given and the pressure is off, the clutch can only open when the torque is off. If there is sufficient tooth play 403 in the circumferential direction, this embodiment can be a sensible solution. The return spring 29; 28; 63 can be designed for a relatively high spring force in favor of a very fast opening process, since in this embodiment it is not decisive for the time of opening or for a torque quantity of the old clutch that is dependent on a friction angle.

The coupling profile can be manufactured very advantageously by machining, for larger series production, in almost all embodiments, also without cutting, using forging technology, sintering technology, sheet metal stamping technology, and extrusion technology. The coupling profile shown in FIG. 3d with interlocking effect or with a negative angle can also be produced without cutting. For this purpose, the corresponding negative angle γ or the interlocking projection 406 is preferably achieved in the cold pressing process by means of an indentation 402 . The bevels of the load-bearing profile surfaces 404 and 405 at an angle γa and bevels or curves 407 on the tooth head ensure an effortless and gentle interlocking during the closing process and during the opening process as a support for quickly returning the coupling ring 408 and the pressure-free piston in neutral position.

The advantage of this clutch toothing lies in the fact that only very low switching forces and therefore very small piston areas and clutch pressures are required. The filling times of the piston chamber during the closing process are correspondingly short. To accelerate the opening process, the piston chamber for the pressure piston 24 is provided with an emptying valve 75 , which has the advantage that after the opening signal, which is triggered spontaneously after the new clutch is closed in this coupling design, the piston chamber can be emptied immediately by the rotational forces, so that after a sufficient drop in torque, the pressure piston or the coupling member 24 can be returned spontaneously via the spring forces 29 .

In a further embodiment according to FIGS . 3b and 3c, the coupling ring 70, 69 is from the piston 61; 62 separately, so that after the opening signal via its own springs 65; 63 the pressure piston 65; 62 can be returned spontaneously. The coupling ring 69; 70 is returned spontaneously via the spring elements 29 and 64 after a sufficient drop in torque.

. In embodiments of Figure 3c, the coupling has separate feedback piston surfaces 217; 219 , which ensure a targeted return of the piston 215 via a corresponding pressure signal after the opening signal within the piston chamber 219 .

The coupling embodiment of Fig. 6 is a friction clutch having conical friction surfaces. On a clutch carrier 90 are rotatably but axially displaceable coupling members 81 and 88 which are axially displaceable against the pressure of the return springs 29 and 28 by means of pressure pistons which can be acted upon by pressure medium and with their conical friction surfaces 82 and 84 against the likewise conically shaped counter friction surfaces 83 and 85 of the coupling members 86 and 87 to be coupled are pressed to close the coupling. The clutch ring 86 or 87 opposite the pressure piston 81 or 88 is supported against an axially fixed thrust washer 91 or 92 mounted on the clutch carrier 90 . When configured as a double clutch, a common thrust washer 91 can advantageously be used for both clutches, as shown in FIG. 6. The coupling can be designed as a multiple coupling, with two or more couplings to be arranged one above the other or next to one another.

The coupling design 80 according to FIG. 6 can absorb a relatively high friction torque or coupling torque depending on the piston surface due to the conical design. The prerequisite is that the friction power is kept low. This is possible in the switching and control device according to the invention in that the switching takes place in the synchronous state and the opening signal for opening the closed clutch only takes place after the torque has dropped by the hydrostatic feedback within both closed clutches by the corresponding correction dimension Z or X plus Y . Due to the almost friction-free shifting, the clutch can be designed for a relatively high surface pressure within the conical friction surfaces and thus for a relatively high torque. In contrast to the known multi-plate friction clutches, the drag torque of this friction clutch is eliminated since the friction surfaces 82, 83 and 85, 84 are deliberately kept separate from one another in the open state. In addition, installation space and costs can be significantly reduced compared to a multi-plate clutch.

When using continuously variable mechanical gears load-dependent speed slip low and their gear elements are very sensitive to slip, this cone coupling is advantageous.  

The coupling version 89 Fig. 6 with conical friction surfaces 85, 84 can also with a driving profile 27 ; Fig. 3a are formed. The advantage of this embodiment according to the invention is based on the fact that larger synchronous deviations in the switching phase can be bridged with a high switching quality, in such a way that synchronous operation can be established via the conical friction surfaces 85, 84 before the clutch teeth are engaged before the synchronizing state is reached. The driving or coupling profile 27 is expediently provided with inclined surfaces or curves 73 which are inclined in the circumferential direction, in order, in particular at higher relative speeds between the two coupling rings 87, 88, to synchronize the conical friction surfaces 85, 84 before engaging after the switching operation has already been initiated or to allow the clutch teeth to mesh. The driving surfaces of the coupling teeth or the coupling profile can be designed as deflection teeth or as straight teeth without an axial deflection effect. Depending on the requirements for shift quality, shift time, clutch torques and others, a largely precise adjustment can be achieved by coordinating piston speed, helix angle of the inclined surfaces or curves 73 , cone angle of the friction surfaces. The piston surface of the pressure piston 88 can be designed to be relatively small in favor of a higher piston speed, since the holding forces are correspondingly small due to the positive coupling type. The coupling profile can be designed in various ways, for example in FIGS . 3a; 6a, 6b and 7 to 7c. With careful design of the cone angle of the friction surfaces 82, 83; 85, 84 and the clutch teeth, taking into account the friction angle, the risk of self-locking is excluded.

Even when the coupling profile is designed as an axial spur toothing 27 , as shown in FIGS . 3 and 3b, an effective synchronization effect can be achieved in that the inclined surfaces 97 and curves 73 , FIG. 3a; 6b are designed to be correspondingly flat in the circumferential direction, so that the clutch teeth can only snap into place after the relative speed between the two clutch rings has been reduced to a correspondingly low level by the synchronizing effect due to the sliding together of the end faces 94 of the clutch teeth 27 . The synchronizing effect can be further increased in that, according to the invention, the axial contact surface for the clutch ring 23 a FIG. 6 b as a conical friction surface 83 a, 98 .

This also applies to the pure friction clutch according to FIG. 6a, in which the clutch ring 86 a has two conical friction surfaces 83 and 83 a, one friction surface being pressed against the axially displaceable clutch ring 81 and the other friction surface against the pressure plate 91 a. The clutch ring 86 a is axially displaceable against a spring element 99 , so that the return of the clutch ring 86 a is possible in the neutral position in order to keep the friction surfaces 83 a, 98 contact-free and drag-free. The double-sided friction surface design according to the friction clutch design Fig. 6c can almost double the clutch torque.

As the circuits are used in continuously variable branching transmissions take place in the synchronous range, the clutch power or friction power or synchronization power relative low, so that the specific surface pressure is relatively high can be used for all friction or synchronization surfaces. Advantageous support is also provided by appropriate Engine speed control within the switching process caused by briefly interrupting the fuel flow or is reduced.

As a synchronizing device in addition to the solutions shown in FIGS. 5 and 6, any known type, for. B. intermediate friction rings can be used as resilient or non-resilient elements, as is already common in known synchronizations.

In contrast to the embodiment according to FIG. 3, the coupling device according to FIGS. 4 and 5 is equipped with a synchronizing device 40 which serves to compensate for any synchronization errors that may occur. This clutch also offers a useful application for multi-step transmissions with or without automatic shifting. Depending on the requirements, the synchronization can be designed as normal synchronization or as lock synchronization. The synchronizer ring 31 is fixed axially in the switching direction on the clutch ring or the piston 24, which is mounted on a clutch carrier 30 in a manner fixed against relative rotation, and is displaceable in the opposite direction against spring elements 38 under spring force. Driver lugs or driver 32 secure the synchronizer ring 31 against rotation relative to the clutch ring or piston 24 . The drivers 32 of the synchronizer ring engage in recesses 37 of the clutch ring or the piston and are secured against rotation via driver surfaces 35, 36 . The driving surfaces 35, 36 are either straight or oblique at a certain angle β in the axial direction. In the event of rotation, an axial force component is so large in accordance with the size of the helix angle β that switching through and intermeshing of the clutch teeth 27 is prevented until synchronous operation is established. When used in a continuously variable power-split transmission, it makes sense to make the angle β smaller so that the locking function cannot be achieved. In this case, this can be so great as a support for the component by appropriately designing the angle β that switching through and intermeshing of the clutch teeth 27 is prevented until synchronous operation is established. When used in a continuously variable power-split transmission, it makes sense to make the angle β smaller so that the locking function cannot be achieved. In this case, it can serve to support the spring force 38 , which can thus be designed for smaller compressive forces. Once the synchronism is reached, the torque of the piston or clutch ring 24 is axially depressed against the pressure of springs 38 for engagement of the clutch teeth 27 up to its stop at the barrier synchronization as a result of slackening. If it is designed as a normal synchronization, as soon as the contact between the two friction surfaces 34 of the clutch ring and the synchronizer ring generates the friction torque or synchronization torque, which is a function of the size of the spring force of the spring elements 38 and the axial component resulting from the smaller helix angle β. If the angle β is equal to zero, the entire axial synchronizing force is brought about via the spring elements 38 . The toothed driving section 27 is advantageously at this coupling embodiment of FIGS. 3, 4 and 5 advantageously as Abweisverzahnung, that is tapered in the axial direction with drive surfaces which cause a torque-dependent pressing of the clutch as soon as the switching pressure is turned off. However, it is also possible to design this driver toothing 27 as a straight toothing or with a relatively small angle, for example below 5 degrees, in order to ensure safe switching off in the no-load state or torque-free state. The piston surface of the pressure piston 24 can accordingly be made smaller. For some applications, it is possible to display the shift position of the clutch by means of a clutch member 30 connected to the piston 24 .

The hydrostat or the stepless converter spontaneously replenishes Closing the new clutch or slave clutch in the opposite direction reversed. That after closing the new clutch The opening signal for opening the old clutch has occurred after a targeted return of the hydrostatic adjustment or of the stepless converter by the dimension Z to the torque of the old to the new clutch within both closed Relocate couplings to the new coupling.

The opening signal for opening the previously engaged clutch that occurred after the area clutch to be newly closed can be generated in different ways or with different devices. The choice of the respective type of device depends on the switching quality required in each case or the type of control device provided, which can be carried out electronically and / or hydraulically and mechanically. According to a relatively simple type of control, it is possible to trigger the opening signal immediately after the sequential clutch is closed, the closing pressure of the sequential clutch serving as an opening signal. The opening time is dependent on the opening travel and the emerging torque of the clutch to be opened within the return time of the pressure piston 24 . This opening time can also be influenced by a targeted pressure reduction of the clutch to be opened, which can not be achieved spontaneously, but by an integrated throttling effect. The throttle effect can also be adjusted depending on suitable operating values, e.g. B. the hydrostatic pressure before the start of switching, signaled, resulting in the measure of the adjustment correction variable X + Y.

Another possibility for the opening signal sees one programmed time after successful closing of the follower clutch before, depending on the application, e.g. B. the size of Can have 0.05 seconds. When using a hydraulic-mechanical A damping valve can be used here to control the triggers the opening signal after a certain valve travel.

Another solution is a permanently programmed one Time size of z. B. 0.1 seconds before, which can be overridden by one or more company sizes, e.g. B. one or both working pressure sizes Δp1 and Δp2 or / and a rate of change the gear ratio, the accelerator pedal or the throttle valve or / and a braking member or brake size. The latter  Possibility makes sense to within the switching phase occurring operational changes, the z. B. by spontaneously changing load conditions, train operation, push operation or through Influence of auxiliary units, for example by a PTO drive occur with machines, lifting devices, etc. can. With a view to a high adjustment speed, in particular within a braking process, is influenced by a brake size, e.g. B. brake force size, hydraulic brake pressure, the switching process is accelerated in that after closing the Follower clutch spontaneously releases the other clutch is granted. A, e.g. B. predefined or preprogrammed Switching time variable can be overridden here by the influence of the brake will. The brake signal is released automatically in the control logic the corresponding signal after the follower clutch is closed. In the case of a hydraulic-mechanical control device, a corresponding one Valve activated. The possible adjustment speed the time is therefore targeted within the switching phase or the hydrostatic variable for opening the clutch determinable.

With regard to sufficient switching reliability, the switching device is designed so that when the old clutch is not opened due to the pressure build-up in the hydrostatic system Hydrostatic adjustment is returned to the torque-free Find the condition of the clutch to be opened. Of another is a shift lock within the non-switchable Adjustment ranges of the stepless converter are provided.

When using friction clutches, multi-plate clutches or in particular The clutch control according to the invention offers cone clutches the advantage that the friction is reduced to almost zero can be.  

A further improvement of the switching device according to the invention provides that within the switching phase before the start of the Area coupling automatically via the control device Torque reduction is triggered by a signal that the Engine control so affected that the fuel supply is brief is reduced or interrupted. The load-dependent adjustment path Correction Z or X, Y is reduced or almost here reduced to zero, which means that the loads during the switching process Links are subject to low requirements. These Solution makes sense e.g. B. in a car to achieve absolute seamless and imperceptible range couplings.

Under the name "synchronous operation, synchronous point" the Understand synchronous range with low speed deviations, which is within a general tolerance range.

In the closing process of the clutches according to the invention Control device designed so that the control pressure is brief is raised until the clutch is closed. In order to is the increased repelling forces of the clutch teeth at Counteracting each other. The control pressures can be in favor the efficiency depends on the load, e.g. B. via a hydrostatic pressure signal, be modulated.

The hydraulic- not described in the description mechanical devices, e.g. B. valves or hydraulic, mechanical, electrical or electronic individual functions are based on known individual elements or on the average specialist known measures.  

The formed according to the invention clutches of Fig. 3 to 6 can be advantageously used in automatically shiftable multi-step transmissions. The switching device is designed so that the new clutch only closes after the old clutch is open and the synchronous state of the coupling members of the new clutch is either manual or automatic by removing the gas during the switching phase or / and by a corresponding synchronizing device, as in FIG. 4 and 5, the synchronous state is established when the clutch is disengaged and the old clutch is open. As is known per se, the engine control or the aforementioned gas removal can be implemented by a specific preprogrammed, in particular electronic control device, the synchronization point for the shift preferably being determined automatically by comparing the speed of two or more transmission members, and then the switching signal as in those already described continuously variable transmission designs for closing the new clutch is issued. In order to avoid possible switching shocks as a result of possible synchronous errors, it is expedient to briefly open the separating clutch, which is alternatively provided as a friction clutch at the transmission input or in the transmission. The synchronizing device shown in the coupling embodiment according to FIGS. 4 and 5 is also suitable for compensating for minor synchronizing errors, as a result of which the mentioned separating clutch can be saved in certain applications.

The clutch designs according to FIGS. 3 to 6 are also particularly suitable in automatic step transmissions which, as is known per se, are designed as double clutch transmissions. The double clutch transmission differs from the usual multi-stage transmission in that at least two powershiftable separating clutches are provided, which alternately take over the power transmission after a gear is selected in the no-load state. Within an already engaged gear, the next gear is preselected in such a way that when the second clutch is disengaged for the next gear, the loose drive train is preferably automatically adjusted and shifted in its speed to the speed of the new clutch via a synchronizing device. With the coupling design 21 according to the invention designed according to FIG. 4, the loose drive train is activated via a synchronizing device 31; 40 synchronized on the new clutch 21 and then closed the clutch in the no-load state. After completing this clutch shift, the shift can now be initiated in the next gear by opening the old and simultaneously closing the new clutch (not shown) designed as a friction clutch.

The switching device according to this invention offers the further advantage when used in the automatic step transmissions described that no shift linkage and mechanical switching devices, as is generally known in step transmissions, are necessary, as a result of which a very compact and vehicle-friendly design can be achieved. The coupling device according to FIGS. 4 and 5 can be produced more cost-effectively and also more economically in terms of space compared to the known mechanically switchable synchronizing switching devices. Furthermore, the modular construction is easier to implement by eliminating the shift linkage and the shift cylinder, and the housing can be manufactured more easily and cost-effectively.

The coupling according to FIGS. 3 to 7 can be designed as a multiple coupling, with at least two couplings being combined one above the other and / or next to one another in a compact and, in particular, short and assembly-friendly design. The pressure oil can be supplied centrally from a shaft in a simple manner.

The coupling profile 76 according to FIG. 3e is equipped according to the invention with high shift teeth 79 and low shift teeth 77 . The advantage of this clutch toothing is based on the fact that very high synchronism deviations can be bridged, which means that a safe clutch shift is still possible at high relative speeds between the two clutch rings 23 and 24 , which is usually no longer guaranteed with shift teeth of the same height, since the axial piston speed is not sufficient to engage the shift teeth, unless a high backlash SV is provided, which leads to the known disadvantages of the clutch backlash in the push / pull behavior of the vehicle. The coupling toothing 76 thus combines the switching advantages of a large backlash SV with the advantages of a backlash-free or largely backlash-free gear toothing. The backlash SV can be designed absolutely free of play with this coupling toothing. Compared to the clutch teeth according to FIG. 7c, which also ensures a high degree of torsional backlash in the switching process, the version 76 has the advantage that there is no play immediately after the switching process and, moreover, that several, usually twice as many teeth can be brought into engagement in favor lower specific load ratios or smaller tooth height, smaller switching paths, lower tooth heights, etc. In the coupling toothing 76, every second tooth 77 is expediently withdrawn by a corresponding dimension S1, so that in the event of a high synchronism error in the switching process, the high tooth 79 does not move into the next tooth gap, but into can grab the next tooth gap by sliding it over the tooth tip 78 of the lower tooth 77 . In the case of a piston travel or switching travel corresponding to the tooth relief by the dimension S1, the inclined driving surfaces 72 of the switching teeth 79 - depending on the design - already meet one another, so that in this state a secure engagement is already possible in this state by sufficient switching pressure. The load-bearing tooth flank 74 can be designed as a repelling or non-repelling driving surface, a higher permanent switching pressure being required for a repelling driving surface and a low permanent switching pressure being required for a non-repelling driving surface in order to keep the clutch in the closed state.

The shift toothing 76 can be used as driving toothing in all of the coupling designs already described. In addition, this coupling design also allows the use of external shift actuation devices, ie, for example, shift devices by means of a shift fork and an external shift cylinder instead of the pressure pistons 81, 88 . In this case, however, it is expedient to design the gear teeth with a non-repellent driving profile 74 . The driving surfaces 74 can also be undercut, that is, with a negative angle γ, in order to rule out possible axial thrusts against the shift actuating device. The inclined surfaces 72 can either be designed with a very small helix angle or straight, non-repellent, driving surfaces or also with driving surfaces, which after initial insertion phases 97 or insertion roundings 73 in the shifting direction cause the coupling teeth to pull together by corresponding, known negatives Tooth flank shape.

The different embodiments of the clutch gears, especially the embodiments that are not repellent Have force components in the switched state are also Can be used advantageously for gearboxes with external, not central pressure piston arranged to the coupling axis. For example a shift cylinder arranged offset to the clutch axis can be advantageous here in connection with a shift fork Actuate coupling ring, if necessary for two clutches, as with step transmissions or also continuously variable powershift transmissions already known.  

The switching device according to FIG. 3f differs from the embodiment according to FIG. 3 in that the tooth coupling ring 23 is not axially rigid, but is resiliently mounted against a pressure plate 91 b and a corresponding spring 50 . The advantage associated with this is that an engagement of the clutch teeth is possible in the shifting process even with larger synchronism deviations. The piston speed of the pressure piston 24 can be lower in this embodiment than in the embodiment according to FIG. 3 in favor of low switching powers or low control oil quantity. In the switching operation so the pressure piston 24 and a relatively high differential speed 23 and 24 of the toothed coupling ring may be at a relatively low piston speed between the two coupling members are pushed back by the spring travel F 23, and then through the spring 50 and the spring rate, the ratchet teeth 76; 71 can be brought into engagement and not, as is necessary with the embodiment of FIG. 3, via the piston speed. The gear teeth are preferably designed as straight teeth, ie as non-repelling teeth. Tooth designs, as shown in FIGS . 3a and 3e, can advantageously also be used, which in the first engagement area allow inclined surfaces 72 to push the tooth coupling ring 23 backwards and, after a sufficient switching path, allow them to snap in via the spring 50 . As a particularly expedient tooth shape is the embodiment of FIG. Verify see 3e as described above, high switching having teeth 79 and low shift teeth 77 and the insertion phase may Zahnschrägungen 97, serve 72 or curves 73 which are variously executed also with regard to tooth wear. The design with high and low teeth, similar to FIG. 3e, can also advantageously be used in the case of straight toothing, the repelling toothed bevels 72 being omitted.

The pressure ring 91 b is expediently supported via a plate spring 50 against a locking ring 49 . In the opposite direction, the pressure ring 91 b and with it the tooth coupling ring 23 in neutral position, for. B. axially fixed by a locking ring 48 .

The clutch according to Fig. 6c illustrates a friction clutch having conical friction surfaces is. This coupling low drag losses connects over the conventional multi-plate clutch, the advantage in the off state despite the high torque transmission capability. Compared to the multi-plate clutch, only little friction or clutch work is possible with this clutch due to the small friction surface. This clutch meets the requirements, in particular, continuously variable power-split transmissions, the shifting sequence of which takes place in the synchronous operation range of the clutch elements to be shifted. Since the friction work in these transmission designs is almost zero, high surface pressures can be accepted, which are possible in the coupling design according to the invention according to FIG. 6c due to the conical coupling elements 334 in a cost-effective and space-saving manner. This clutch can also advantageously be used as a clutch for multi-step transmissions, in particular with an automatic shifting device. By eliminating shift linkage, shift forks, shift cylinders, very compact and space-saving gearbox designs are possible.

The coupling according to Fig. 6c consists of a mounted on a clutch carrier 331 rotationally fixed but axially displaceable with a conical friction surface 337 provided coupling ring 345th A further clutch body 341 is provided on the same clutch carrier 331 with a conical friction surface 339 running parallel to the other clutch ring 345 . Between these two conical friction surfaces of the two clutch rings rotatably mounted on the clutch carrier, a clutch body 334 , which is designed with conical friction surfaces 338 and 336 , is arranged, which is in rotary connection with the second clutch half or with the clutch member 19 to be coupled. The coupling body 334 of the second coupling half is axially movable. When the clutch is closed, the clutch body 334 is pressed between the two conical friction surfaces 339 and 337 of the two clutch bodies 345 and 341 . The coupling ring 345 is preferably acted upon by a pressure medium in the axial direction via the piston 332 . By spring elements 343; 344 the two conical friction surfaces are kept apart in the neutral position. The friction body 334 of the second coupling half is expediently designed as a stamped sheet metal part. It has, depending on the embodiment and the subsequent gear member 19 , driver tabs 347 , which are preferably formed in a stamped and in any shape adapted so that a large driving surface compared to the other gear member 19 is possible. The driver tabs 347 are designed here in a curved shape along a driving toothing so that a low or sufficient surface pressure and a largely resistance-free axial displacement and adaptation to the relevant switching elements 341, 345 is possible. The pressure body 341 with a conical friction surface 339 can also be produced in a cost-effective manner as a stamped sheet metal part. As a common component, it can also form the pressure body 342 of a further clutch 340 . Here, the common pressure body 341, 342 is mounted on the common coupling carrier in a very advantageous manner so that only a single locking ring 352 is necessary for axial fixation, the rotary connection taking place via recesses 349 of the coupling carrier and 348 of the pressure body. The second clutch 340 is in this case designed as an internal clutch, the friction ring 335 having its second clutch half connected to an internal shaft or gear 353 via a non-rotatable but axially displaceable connecting member 354 .

. The cone clutch of Figure 6c can also with two or more friction elements 334; 335 (not shown) are formed, the pressure body 341; 342 is designed to be axially movable at least to a limited extent and a further pressure plate which is axially and non-rotatably connected to the coupling carrier is provided as the end pressure body. Here, the number of friction surfaces is increased by the number of friction bodies 234; 335 multiplied and the friction torque increased accordingly at the same switching pressures.

The essential advantage of this conical friction clutch according to FIG. 6c is, as already mentioned above, that very high torques are achieved by the cone effect, which can be increased further by doubling or increasing the number of friction surfaces. In addition, there is the advantage that the drag torques compared to the multi-plate clutch are almost zero, since the centrifugal force does not guide the oil between the friction surfaces, but on the outer conical friction surface in such a way that the counter friction surface remains oil-free and thus free of drag loss.

For lubrication of the friction surfaces are impressions in the as Embossed sheet metal parts or also as sintered parts Coupling elements provided that are not in the drawings are shown. Because the friction work is very low or almost zero is, it is also not necessary to use this coupling in continuously variable branching gears with special to provide costly friction linings. In most There is no need for post-treatment of the friction surfaces on the coupling elements designed as stamped sheet metal parts.

The individual components of this coupling according to FIG. 6c, in particular the components provided with the conical friction surfaces, are particularly suitable for the use of modern chipless production, such as forging technology, sintering technology, sheet metal technology.

The shift cone clutch according to FIG. 6d is particularly suitable for gear members 368 to be coupled to the gear housing. In this clutch design, for example, a ring gear 368 of a planetary gear is coupled to the gear housing. The gear housing simultaneously represents the clutch carrier 360 of the cone clutch. The gear member or ring gear 368 has one or two opposing conical friction surfaces 364 and 366 which interact with corresponding conical friction surfaces 365 on the housing and 361 of a clutch ring 363 which can be acted upon by a pressure piston 369 . The double cones 364 and 366 have the advantage of doubling the frictional torque in favor of smaller piston forces and their associated switching pressures and switching powers. Coupling ring 363 and piston 369 are returned to the neutral position by a spring device, not shown, in the switched-off state.

With regard to high shift quality and low friction, when using the friction clutch in stepless power split transmissions, an improved shift sequence can be determined in such a way that after synchronous operation has taken place, the new clutch is closed and the old clutch is subsequently opened. Within both closed clutches, the load-dependent speed slip of the stepless converter is compensated for by adapting to the load condition in the new switching range, the stepless converter being corrected in its variable by the dimension X and Y or Z as in when the old clutch is still closed shown FIGS. 2 and 2a and described in more detail already with the use of toothed couplings and other friction clutches. The correction variable Z can be obtained from a pressure signal, e.g. B. the hydrostatic converter or other operating variables, as already described, can be defined or programmed. The opening signal after adjustment correction Z expediently takes place in this clutch after the old clutch has been completely relieved, that is to say after the new clutch has taken over the full torque, which corresponds to shift point SK1 (see FIGS. 2 and 2a). The adjustment correction dimension Y is generally greater than the correction dimension X. Based on constant load conditions or tractive force relationships, the relation X corresponds to the hydrostatic pressure before switching within the old switching range and Y to the hydrostatic pressure at the beginning of the new switching range. Thus approximately Y = X · Δp2 / Δp1. Their size ratios depend on the gearbox design and essentially on their area size ratios. Correction dimension X is determined from the actual manipulated variable αSK2 for synchronous operation of the clutch elements of the new clutch and the theoretical manipulated variable αS according to the no-load condition. The synchronization signal is preferably determined from a speed comparison, as is known and described at the beginning. Correction variable X represents the difference between the actual manipulated variable αSK2 of the stepless converter and the theoretical manipulated variable αS. The corrective measure Z = X + Y, which determines the opening signal, can thus be determined from the manipulated variable difference α SK2 and αS or by the hydrostatic pressure before the start of the shift and specified accordingly or be programmed. This applies to both the tooth clutch and friction clutches. In the case of toothed clutches, an opening signal which starts earlier, ie a corresponding reduction in the correction dimension Y, can be taken into account, taking into account the opening path for disengaging the gear teeth and the opening time associated therewith.

Load fluctuations within the switching phase, d. H. after already initiated shifting require a change in Correction dimension Z, since the clutch torque TK2 also compared TK1 changes before and after the range switching accordingly. An optimal adaptation of these operationally changing conditions  is within the switching phase by a torque signal from the Given the torque of the old clutch to be opened, which Opening signal from the torque-free state of the old clutch is signaled. An approximate adjustment can also be made, as in the Main application or already described above by the beginning working pressure Δp2 for the new switching range B2 be. When using a tooth coupling is in most Use cases the opening period, which is between the Switch point S and the actually open state up to SK1 sufficient to achieve a good shift quality. When using a friction clutch can also have a slight delay effect in the clutch opening process, e.g. B. by a targeted Throttle effect can be achieved in the clutch switching valve by the ideal switch point SK1 to reach or almost come.

The pre-programmed correction dimension Y determined from X can be determined by external operating influences such as accelerator pedal changes, brake actuation or external load changes, influenced or overridden will. The corresponding signals from the accelerator pedal or Signal to change the fuel supply, brake signal, load signal, e.g. B. from the hydrostatic pressure and / or other signals that the continuous course within the switching phase influence constant load condition.

If the gas is removed during the switching phase, i.e. H. at already closed new clutch is z. B. the opening signal for the old coupling immediately or after a correspondingly smaller correction triggered, with other described correction functions can be overridden or adjusted accordingly by Actuation signal or an from the accelerator pedal actuation resulting signal. Throttle removal can alternatively be done in control logic programmed and automated, whereby after closed new clutch, the fuel supply is interrupted or reduced  is by corresponding control signal. The correction size is thus different according to the size of the accelerator pedal movement can be dosed by simultaneously influencing specified ones Correction function.

To increase the switching reliability when the switching process is initiated, the reversal of the direction of adjustment for the stepless converter can be carried out according to a certain characteristic, which is particularly advantageous for gear clutches. For this purpose it is provided according to the invention that the change in the direction of adjustment does not occur spontaneously, but instead has a smooth, gradual speed profile, so that even when a low piston speed of the pressure piston 24 is taken into account, a secure interlocking of the gear teeth is ensured. This can be brought about by a correspondingly provided throttling effect or delay effect in the adjusting device of the stepless converter.

Another possibility for switching protection sees between the two high pressure lines of a hydrostatic transmission provided bypass valve before that within the switching phase is opened briefly, with the bypass line with or without an additional throttle or / and a pressure relief valve which is set to a relatively low pressure, so that higher differential speeds or Synchronous errors, especially when shifting with a gear clutch can be bridged without damage.

With continuously variable power split transmissions, such as. For example, shown in FIG. 8, the power accumulated in a summation planetary gear 251 flows alternately via two or more output shafts 240, 242 or 243 and is transmitted via intermediate elements 245; 250; 249 with one or more closed further clutches 236, 237; 246; 248 transferred to the output shaft 233 . In certain shift ranges, depending on the clutch engaged, one or more of the aforementioned transmission members rotate in the no-load state, ie in the torque-free state. In order to avoid an uncontrolled co-rotation of these transmission members in view of the fact that synchronous operation of the members to be coupled is ensured for the switching operation, z. B. one or both of the coupling members 236, 237 held on the housing. This is done by a locking device, not shown, which can be independent of the clutch 236, 237 designed as a toothed clutch. In the illustrated transmission scheme of Fig. 8, the clutch 237 is closed in first and second forward drive range and the clutch 236 in the first and second reverse driving range. When the engine is started or when the engine is idling, ie when the shift is in the neutral position, the respective clutch elements 250, 249 would rotate without load due to internal friction or drag resistances and possibly lead to damage to the shift teeth when the respective clutch is closed. The locking device, not shown, can be a device which is dependent on or independent of the clutch, and which, by spring force, transmits the relevant transmission element 250 or 237 in the form of a braking device or a form-fitting device, e.g. B. locked as a toothed element, this engages in a corresponding profile of the aforementioned gear members, or as a friction element, the gear member - ring gear 249 , web 250 - braked or locked by spring force when starting the engine and / or in neutral position with respect to the gear housing. After starting the engine or after preselecting the direction of travel, the corresponding transmission link is hydraulically unlocked or released. The locking device can be a simple, separate component that presses or locks onto the relevant transmission link from the outside.

The invention further sees the combination of a friction clutch 234; 235 with a tooth coupling 236, 237; 246 before. In the switching phase, after triggering the synchronous signal or the switching signal, the tooth clutch 246 or 237 is switched first and then the relevant friction clutch 235 or 234 . The torque load in spite of a relatively high synchronous error of the links of the tooth clutch to be closed is very low in this shift or clutch combination, since the relevant link 245 only has a low torque at the start of the shift or in an almost torque-free state with the relevant output shaft 240, 242 of the summation planetary gear 251 co-rotates and the said synchronous error is compensated for by possible slipping of the friction clutch in question. The intermediate gear member, shaft 245 , is, as mentioned, torque-free with one of the two output shafts 240, 243 , ie coupled with a low shifting pressure, or the friction clutch in question is opened spontaneously at the start of shifting before the toothed clutch engages and then, e.g. B. immediately closed again by the closing signal via the oil pressure of the closed tooth coupling. The transmission intermediate member 245 involved in the power transmission in the new shift range is carried without load in the old shift range without torque with the clutch 235 in question and the relevant output shaft 242 or 240 of the summation planetary gear 251 in order to ensure synchronous operation of the clutch elements when the new clutch (tooth clutch 246 ) is closed and to avoid uncontrolled rotational behavior of the intermediate gear link 245 . The area coupling 234 or 235 is alternatively designed as a cone coupling, as shown in FIG. 6c.

The invention provides a safety device in the event that it is not possible to close a tooth clutch due to a control function failure or excessive synchronism deviation in the switching phase. This safety device is based on the fact that the setpoint of the gear ratio in the control device, in particular the electronic device, is compared with the setpoint of the variable of the stepless converter by comparing two or more rotating gear members. If the adjustment value of the continuously variable transmission 232 does not match the setpoint value, a safety signal is automatically triggered in the event of a deviation to an admissible extent, which causes one or more of the clutches to open in order to put the vehicle in a torque-free state and / or that the adjustment member of the stepless converter is automatically adjusted so that the clutch can be closed again at the given gear ratio.

Another safety device according to the invention provides that z. B. with excessive deceleration or braking of the transmission and insufficient adjustment speed of the stepless Regulation open one or more of the clutches and one Establish the no-load condition of the transmission links, either after this process the vehicle is braked to a standstill must or that after changing this operating state an automatic readjustment a match of the Adjustment variable of the adjustment element of the stepless converter with the Actual value of the gear ratio is sought, after which the Established synchronization point of the area coupling concerned and this clutch is closed automatically.

A more detailed description of the transmission system according to FIG. 8 is contained in DE 39 29 209 A1.

Claims (14)

1.Switching device, in particular for an automatically switchable motor vehicle transmission, which consists of at least two switchable clutches and at least one clutch is designed as a toothed clutch, a toothed coupling ring ( 24 ) of the toothed clutch being provided by a switching device against a counter-rotating toothed clutch ring ( 23 ) Intermeshing of the two clutch rings is axially displaceable, characterized in that the counter-directed clutch ring ( 23 ) is resiliently slidable in the axial direction and that the gear shifting takes place in the synchronous range of the shifting elements to be coupled ( Fig. 3f). 2. Switching device, in particular for automatically switchable motor vehicle transmissions, for switching two or more shift ranges, in which the shift takes place predominantly in the synchronous range of the clutch elements to be closed, characterized in that the clutch ( 330; 340 ) is designed as a cone clutch, such that one coupling half has coupling elements ( 341 and 345 ) which can be displaced relative to one another by a pressure medium and have parallel conical friction surfaces ( 339 and 337 ) and the other coupling half has one or more coupling members ( 334 ) provided between these conical friction surfaces and provided with corresponding conical friction surfaces ( 338 and 336 ) ; 335 ) ( Fig. 6c). 3.Switching device for a continuously variable branching transmission with two or more shift ranges, which comprises a continuously variable converter and a summing planetary gear with at least two output shafts ( 240; 242 ) which can be shifted with a downstream shift transmission ( 247 ), the downstream transmission shifting clutches ( 237; 236, 246 ), which alternately establish a drive connection via one of the output shafts of the summation planetary gear with the output shaft ( 233 ), characterized in that at least one of the output shafts of the summation planetary gear via a friction clutch ( 234; 235 ) and another downstream clutch ( 236) ; 237, 246 ) establishes a drive connection with the output shaft ( 233 ), the gear clutch and then the friction clutch being closed first for shifting a new shift range and that this friction clutch within the old shift range has no load on it m in the new shift range torque-transmitting gear member ( 234 ) is coupled so that when the tooth clutch closes, an at least initial slipping of the friction clutch with low torque is possible, or that the friction clutch opens briefly before the tooth clutch closes ( Fig. 8). 4. Switching device according to claim 1, characterized in that the coupling teeth as deflection teeth (according to FIGS . 7 to 7c) or as non-repellent coupling teeth or as coupling teeth with repelling driving surfaces ( 72 ) and non-repelling load-bearing driving surfaces ( 74 according to FIGS . 3a and 3e) or / and is designed as a coupling toothing with high and low teeth ( 79, 77 , Fig. 3e). 5. Switching device according to claim 1 and 4, characterized in that the opposing tooth coupling ring ( 23 ) with a straight contact surface or conical contact surface (not shown) is supported on its pressure plate ( 91 ). 6. Switching device according to claim 2, characterized in that the second coupling half has two or more conical friction bodies ( 334 ) (not shown) which cooperate with a corresponding number of friction elements ( 341 ) of the first coupling half. 7. Switching device according to claim 2 and 6, characterized in that the pressure body ( 341, 342 ) provided with one or more conical friction surfaces forms a common component for two clutches ( 330, 340 ). 8. Switching device according to claim 2, 6 and 7, characterized in that the with at least conical friction surfaces provided coupling elements are partly produced without cutting and molded without cutting Have recesses for the lubricating oil supply (not shown). 9. Switching device according to claim 2 and 6 to 8, characterized in that two clutches form a switching function unit in such a way that after Synchronous operation closed the new clutch and then the old clutch is opened and that when used in one stepless transmission or stepless power split Gearbox load-dependent speed slip within both closed clutches by an appropriate adjustment correction a stepless converter using an appropriate correction measure (X + Y or Z) is adjusted before opening the old clutch.   10. Switching device according to claim 3, characterized in that the one Toothed clutch friction clutch in operative connection Multi-plate clutch or a cone clutch. 11. Switching device for a continuously variable power-split transmission with a continuously variable converter ( 232 ) and a multi-shaft summation planetary gear, which has switching gear clutches for switching several shift ranges, characterized in that a gear member to be coupled for the forward driving range and for the reverse driving range is automatically locked by a locking device when the vehicle is at a standstill or is held in the neutral switching position and / or after starting the engine and that these transmission members are preferably members which can be connected to the housing or are dependent on the coupling elements and / or are independent devices which prevent uncontrolled rotation of one or more transmission members in the no-load state . 12. Switching device according to the preamble of claim 3, characterized in that a control device is provided, which is based on a speed comparison of different Transmission links determine the gear ratio and from this the actual value and the setpoint of the variable of the stepless Converters checked and that in the event of a deviation from a permissible Value a safety signal is triggered, the one or opens multiple switching clutches and / or that through the control and Control device automatically adjusts the adjustment variable, to a relevant clutch again in synchronous operation or in Close synchronous running range. 13. Switching device according to the preamble of claim 2, characterized in that the adjustment Correction variable X is the difference between actual  Manipulated variable αSK2 and the theoretical manipulated variable αS and that the size ratio of the correction quantities X to Y den theoretical hydrostatic pressure ratios Δp1 before shift start corresponds to Δp2 after the range switching. 14. Switching device according to claim 13 and another of the Claims 1 to 13, characterized in that the correction dimension Y can be overridden or changed by Change of operating state during the switching phase by Change in the accelerator pedal signal and / or influence of the brake pedal Signal and / or change in the external load state.